Centripetal turbine



Feb. 1, 1966 R. BIRMANN 3,232,530

CENTRIPETAL TURBINE Filed July 18, 1963 4 Sheets-Sheet 2 INVENTOR RUDOLPH BIRMANN BY Feb. l, 1966 R. B|RMANN 3,232,580

CENTRIPETAL TURBINE Filed July 18, 1963 4 Sheets-Sheet 3 *6 @IEW/ F l G. 5.

INVENTOR. RUDOLPH BIRMANN ATTORNEYS Feb. 1, 1966 R. BIRMANN 3,232,580

CENTRIPETAL TURBINE Filed July 18, 1963 4 Sheets-Sheet 4.

INVENTOR.

. RUDOLPH BIRMANN .LJ/f' fm; ATT YS United States Patent O 3,232,580 CENTRIPETAL TURBINE Rudolph Birmania, Highland Farm, R.-D.- 1, Newtown, Pa. Filed July 18, 1963, Ser. No. 295,874 7 Claims. (Cl. 253-39) This application is in part a continuation of my application Serial Number 85,851, iiled January 30, 1961, as a division of my application SerialNumber 825,773, filed July 8, 1959, now Patent 3,059,415, issued October 23, 1962.

This invention relates to centripetal turbines and particularly to blade constructions therefor.

In accordance with the invention there is provided a novel configuration of centripetal turbine blading which prevents the separation of the flow that occurs on the leading side of conventional centripetal turbine blades, and which lconsequently results in improved turbine eiciency. Additionally, separation and shock losses at the turbine blade inlet edges are also prevented, even under conditions of operation wherein the angle of the ow relative to the blades differs greatlyfrom the blade inlet angles, so that the range of operation over which high efficiency is achieved is greatly extended.

l'It is, therefore, one of the objects of the present 4invention to improve substantially the turbine elciency.

Another object is to maintain the eiciency high over a very wide range of L11/co ratios; i.e., the ratio between the blade tip speed u1 and the spouting velocity C0 that corresponds to the total turbine enthalpy drop. In accordance with the invention the eiiiciency is substantially maximum and uniform through a range of this ratio extending from less than 0.5 to more than 0.9, with this maximum efficiency at 83% to 85%. In comparison the best results obtained from prior turbines for comparable use, size and type (centripetal) have shown eiciencies of 78 to 80% through a range of the ratio between 0.6 and 0.8, the eiciency in this last case being approximately 70% at ratio values of 0.5 and 0.9.

In accordance with the invention there is also achieved a characteristic of operation which is particularly desirable for some turbine uses, for example, the use in turbochargers. This is that the equivalent orifice area decreases as a function of increasing :l1/co ratio. The term equivalent orifice area is here used in its usual sense of referring to the area of an orifice which, if substituted for the turbine, would present the same characteristics to the driving gases. The characteristic means that as C decreases and, therefore, the ratio increases, u1 being approximately constant over a useful working range, the turbine presents to the supply an increasing effect of damming up the gases, whereby most effective use is made of the available energy. In prior centripetal turbines the corresponding characteristic is that of increase of equivalent orifice area with increase of the ratio.

Still another object is to provide structurally much more rugged blades which are not easily damaged by foreign objects that may become entrained in the gas flow and which are not subject to blade vibration and the possibility of failure therefrom.

A further object is to prevent foreign particles that enter the rotating blade passages from being flung radially outward only to reenter again and again, thereby destroying the blade inlet portions.

Still another object is to minimize the total wetted blade surface for the purpose of improving the efficiency of the turbine and of making its performance less susceptible to Reynolds number effects and, at the same time, of making it a relatively easy problem to maintain low blade metal temperatures by internally cooling the biades.

The attainment of the foregoing and other objects of ICC the invention will become apparent from the following description, read in conjunction with the accompanying drawings, in which:

FIGURE l is an axial sectional view showing the turbine in association with a housing and other parts of a turbocharging unit, the blades being shown in circumferential projection;

FIGURE 2 is a composite view showing the turbine rotor, the upper portion of this figure being an elevation of the right-hand or discharge side thereof looking radially towards the left in FIGURE 1, the lower left-hand portion A of the figure showing a radial section taken on the plane indi-cated at 2A-2A in FIGURES land 3, and the lower right-hand-portion B of the figure showing a radial section taken on the plane indicated at 2B-2B in FIGURES 1 and 3;

FIGURE 3 is a View, developed into a plane, of the projection of the turbine rotor on a cylinder coaxial therewith;

FIGURE 4 is a plot illustrating variations of orthogonal passage area-with mean line distance of ow;

FIGURE 5 is an axial section of a modified rotor having a zero axial component of inlet flow and provided in accordance with the invention, the housing being indicated only in outline, and the blades being shown in circumferential projection;

FIGURE 6 is a development into a plane of a section taken on the cylindrical surface indicated at 6-6 in FIGURE 5;

FIGURE 7 is a development into a plane of a section taken on the lcylindrical `surface indicated at 7 7 in FIGURE 5;

FIGURE 8 is a composite view showing the turbine rotor, `the upper portion of this figure being an elevation of the right-hand or discharge side thereof looking axially towards the left in FIGURE 5, the portion C of the figure showing a radial section taken on the plane indicated at 8C-8C in FIGURES 5, 6 and 7, the portion D of the figure showing a radial section taken on the plane indicated at 8D-8D in FIGURES 5, 6 and 7, and the portion E of the figure showing a radial section taken on the plane indicated at SE-SE in FIGURES 5, 6 and 7; and

FIGURE 9 is an lelevation of the left-hand side of the rotor looking axially towards the right in FIGURE 5.

The turbine forming the subject matter of this application is of general utility, but it is particularly advantageous in turbochargers as fully described in said Patent 3,059,415. In the present application only so much of the associated construction details will be described as are necessary for an adequate understanding of the invention.

A rigid, stress-sustaining portion of a housing is formed by castings, one of which is indicated at 2 and provides a left-hand bearing (not shown). The other is indicated tat 6, and has formed as part thereof the right-hand bearing housing 8. This part 6 is primarily in the form of a spider which is secured to the part 2 by bolts 10. A sheet metal cylinder 12 has an annular iiange 16 clamped between the parts at 2 and 6 and supports a sheet metal annular member 14 which is held in place by screws 18; The cylinder 12 is shrunk over the concentric ceramic filler 20 within which is retained (by a retaining pin not shown) ceramic filler 22, the fillers, of refractory material, being separate merely to facilitate assembly. The shrinking of the cylinder 12 on the filler 20 results in compression. The ceramic filler will withstand high compression but no substantial tension, and the arrangement insures tightness despite the large differences in coefcients of thermal expansion of the cylinder and ller, any possibility of tension stresses in the ceramic being eliminated by its prestressing in compression. The filler 20 provides the vaneless vortex space 24 for the feed of driving gases from the engine or other source of gases to the turbine blad-ing, there being provided nozzles (not shown) for directing the driving gases into the space 24 at high veloci-ty in a tangential direction. The turbine driving gases may originate as intermittent pulses from an engine, but by providing a substantial radial region in the space 24 between the nozzles and the inlet edges of the turbine blades, the gases spiral inwardly and vortex ow automatically occurs to minimize pulsations at the inlet edges of the turbine blades.

The ller 20 `and the member 14 provide between them the annular vaneless diffuser 32 for the gases exhausted from the turbine. rlhe diffuser provides a radially outward component of flow to the gases and discharges them through a suitable conduit which conveys them to a desired exhaust point.

The turbine rotor 56 carries the turbine blades 57 which will be described in detail. The rotor is integrally formed with a shaft 58. The mounting of the shaft arrangement and hubs is disclosed in detail in said Patent 3,059,415 and form no part of the invention as herein claimed.

Considering the improved turbine design, reference may first be made to some general considerations. The efficiency of centripetal turbines should be theoretically much higher than they have been in practice. Loss coeiiicients applicable to the How through the turbine blading have been unusually high. M-ore conservative blade loading, design of blade shapes to secure iiow balance and increase of the degree of reaction to secure highly accelerated iiow (known to decrease ow losses) have only been moderately successful in bettering the efficiency. Blade vibration suggests that losses may be due to highly disturbed nature of the ow through the blade passages due to boundary layer phenomena. Without here reviewing the evidence therefor, the conclusion was reached that the losses were largely due to actual flow separation from the leading side of a blade giving rise to a zone of turbulence (stagnant relative to the blade) which restricted the channel through which smooth ow could occur. Further, it appeared that this was due to centrifugal forces acting on the boundary layers which, rotating with the blades, may be viewed as bodies subject to such forces. The boundary layers were thus subject to pressure gradients negative with respect to the direction of flow, and such negative pressure gradients result in rapid growth of boundary layers. The overall phenomena involved rapid recurrence of transients accounting for the noted vibrations.

The adverse pressure gradient within centripetal turbine blade passages is proportional to the square of the change of the circumferential velocity which, typically, at blade tips may be 157() feet per second. To suppress the growth of the boundary layers it is necessary to offset the adverse negative pressure gradient caused by centrifugal forces by a positive pressure gradient which is derived from accelerating the relative velocity of iiow within th-e blade passages. To do this the areas orthogonal to the flow should be so adjusted that the relative velocity increases with decreasing radius, and consequently the orthogonal passage area should decrease continuously from inlet to outlet in a special pattern in relationship to the change of radius. The cross-sectional passage area should change in accordance with the above so that plotting, as in FIG- URE 4, ortlogonal area against a mean line distance of flow the resulting curve M should be concave upwardly rather than concave downwardly (as indicated at N) as at present in conventional centripetal turbines, this being trueeven if theterminal cross-sectional areas at inlet and outlet were the same for both plots, i.e., the total reaction were the same in both instances. In FIGURE 4 the inlet is indicated at I and the outlet at O, the plot being in terms of percentage of theorthogonal area at the inlet which is taken as 100%.

In the case of the old type of blading the heat drop of the driving fluid assigned to expansion in the blading is i expended mostly at the discharge, whereas it should be mostrapidly expended in the initial portion of the blading` An attempt to change the cross-sectional area available to flow in accordance with the above by changing the blade tip contour is readily found to be impractical on the basis of stresses which would arise, meridional disturbances, manufacturing diiculties, and the like. An alternative solution is, accordingly, provided in accordance with the invention by making the blades sufficiently thick so that the space which is not desired for iiow passages is effectively filled.

In accordance with the present invention, a blade construction is utilized which will be clear from consideration of the drawings. Assuming, to start, given hub and eripheral meridional blade contours and given circumferential projections of inlet and outlet edges, for example, as illustrated in FIGURE 1 (these being chosen in accordance with conventional considerations of the centripetal action desired, aerodynamic loading, etc), a flow passage may have its mean surface determined in accordance with conventional computations, taking into account in the design a variation of cross-section of the flow passage normal (orthogonal) to the flow to provide a velocity distribution in accordance with the foregoing, i.e., involving a pressure gradient due to reaction which at each point exceeds the negative pressure gradient due to centrifugal effects. It will then be found that the passage walls 122 and 24 will be as illustrated particularly in the development of the circumferential projection forming FIGURE 3. If the walls of adjacent passages are then combined to form blades, i.e. the leading wall of one passage with. the trailing wall of the next prece-ding, there is found the fortunate circumstance that the inlet edges are circumferentially thick and by rounding become highly bulbous as indicated at 120, the sides of the resulting blades at 122 and 124 converging to a discharge edge at i126 which may be rounded to provide for a smooth ow pattern. To provide a light construction the interior of each blade thus arrived at may be opened as at 123.

It is further found that with negligible departure from theoretical shape the thin walls extending from the portion of the bulbous inlet edge 120 toward the location of discharge may be quite closely radial. The blades as a whole being radial and therefore not subject to bending stresses under conditions of high speed operation. The sections shown in FIGURE 2 clearly show the radial extents of the blades, as does also the developed circumferential projection, FIGURE 3.

Summarizing the foregoing, a passage is first designed having its orthogonal cross-sections chosen to provide the specified velocity distribution; assuming two such passages' 'adjacent to each other with their exit openings approximating contact, the material between them becomes a theoretical blade; the inlet portion of this blade will then be found to be rounded, providing a bulbous entrance edge, and by slight changes in the passage boundaries the walls may be made approximately radial; finally, removal of interior material results in the final blade shape, the remaining material thickness being chosen to meet centrifugal stress considerations.

In FIGURE 3 the developed radial proiections of the streamlines of ow at the hub are indicated at S. The traces of the areas orthogonal to the flow are indicated at T, these traces being those existing at the hub. T he cornplete areas are, of course, of surfaces orthogonal to all lstreamlines through the passages and the surfaces have resulting curvatures. These orthogonal areas are those to which the plot of FIGURE 4 relates.

The bulbous inlet edges of the blades constitute archlike constructions convex with respect to the axis of rotation, which may be so shaped as to conform with surfaces; which involve `the metal adjacent to these inlet edges being substantially only in tension and not subject to bend-v ing stresses due to Centrifugalforces. The open spaces;

128 within the blades are desirably in communication with the back of the hub through individual passages 13) to receive air from a compressor about the shaft 58. In the case of turbochargers for diesel engines, cooling is not of great importance; but if a turbine is to be operated at high temperatures special provisions may be made for introducing an adequate ow of air for cooling purposes through the passages 130 into the openings 12S, to be discharged either at the periphery or otherwise into the gases flowing through the turbine blading.

The described turbine construction has Various advantages, some of which have already been mentioned, as follows:

Owing to the strong acceleration of the llow immediately upon entering the rotating blade passages an excessive rate of boundary layer growth and separation of the ow from the blade surfaces are prevented. As discussed above, this is an action peculiar to centripetal turbines in which the ladverse effects of a pressure gradient set up by radially inward iiow against centrifugal force must be overcome.

While in ordinary turbine blading moderate aerodynamic lift coeicients are achieved by increasing the nurnber of blades, the securing of moderate lift coeilicients in the present instance is secured by reason of the fact that the lift coetiicient decreases inversely proportionally to the increase of number of blades but also inversely proportic-nally to the square of the increase of the velocity of ow relative to the blades. Accordingly, comparing the present construction having, for example, nine blades with another construction of similar dimensions but having eighteen thin blades, tlie lift coeticient is doubled by using one-half the number of blades, but by reason of the doubling of relative velocity through the blade passages the lift coelicient is reduced by a factor of four. The end result is that the aerodynamic lift coetlicient for the nine new thick blades is onehalf that for the eighteen old thin blades. This means that ti e new turbine wheel is much more conservatively designed so far as lift coeiicients are concerned.

By providing fewer (but thicker) blades in accordance with the invention, the total friction surface is drastically reduced and, moreover, the Reynolds number vapplicable to the flow passages is substantially increased. Both ot" these factors additionally lead to reduction of flow losses and to corresponding improvement in turbine efficiency. While this is true for centripetal turbines of all sizes, the improvement is particularly significant in small turbines where friction losses and other Reynolds number effects constitute such a large percentage of the total losses that in the past it was impossible to achieve acceptable efficiency in small turbines. For example, `a highly efficient turbine produced in accordance with the invention had a mean inlet diameter of three inches and had a normal operating speed ranging from 100,000 to 120,000 rpm.

Any turbine blade derives its driving `force from the pressure dierence that exists across the blade. rlfhis pressure diterence causes a leakage iiow between the tip of each blade and the stationary shroud surface. The leakage flow constitutes a serious loss, partially because it reduces the blade driving force and partially because .it disturbs the main ow.` The .ew blades involve sealing against such How by means of the labyrinth etlect particularly resulting from the provision of a large expansion space (QS) between the two labyrinth throttling points: the outer edges of the blade walls at F122 and 24. ln conventional centripe-tal turbines the angle of the relative iiow entering the blade passages must accurately coingide with the entrance blade anle but this is hardly ever the case, and at best only at one single operating 1 peint, because the blades. @jf the conventional eeitripetel turbine are always very thin exposing an almost sharp edge to the ow, so that even a slight deviation of the relative ow angle from the blade angle results in shock and separation losses. In the case of the novel blading disclosed herein the entrance edges turn out to be bulbous by virtue of the described design of the ow passages.V

The flow around the inlet edges is very highly accelerated. These two factors combine to prevent ow separation at the inlet and the losses connected therewith even if the relative flow angle at the inlet differs greatly from the blade or passage angle. The range, accordingly, over which eilicient operation can be maintained is greatly extended so that the new centripetal turbine may handle many problems which cannot be solved with the conventional type of turbine. For example, this is true when `the turbine drives a variable-pressure-ratio turbocharger compressor as described more fully in said Patent 3,059,- 415. In such a case the turbine must operate with a ttl/e0 ratio both lower and higher than thatl (for example 0.55) for which it is designed and the reaction and the blade angles chosen to result in optimum eliciency. Lower and higher L11/C0 ratios both give rise in conventional turbines to very poor turbine eiciencies. In the case of lower L11/co ratio (0.40 for example) the efficiency drops olf very greatly because of shock on the trailing side of the blades at their inlets and the blade inlet angle not coinciding with the relative flow angle triggering oli complete ilow separation which is primarily due to insufficient acceleration on the leading side. At high 111/00 ratio (1.00, for example) energy dissipating shock occurs on the leading side, separation on the trailing side, and due to insuicient reaction, the gases are discharged while still retaining a large whirl component, all of which results in a precipitous dropping od of the efficiency. This may even extend into a condition in which braking occurs, this condition arising under en-4 gine idling conditions.

During such engine idling operation the energy available to the turbine is a minimum and because of the high tf1/C0 the kinetic energy at the outlet of the turbine blades as percentage of this energy available to the tun bine is a maximum.` Under these conditions, so as to achieve maximum possible output from the turbine, the radial exhaust dilluser 32 becomes of particular importance. It converts the high (percentagewise) kinetic energy at the turbine blade discharge into a pressure rise which results in the expansion within the turbine blade passages to be carried to a lower back pressure (even to a sub-atmospheric pressure) and consequently in the at; tainrnent of increased output.

As pointed out in said Patent 3,059,415, the described turbine, when used in a turboconjipressor, in connection with an engine, has special advantages which need not be described herein,

A turbine having diagonal flow as illustrated is para ticularly advantageous since by such an arrangement the inlet angle may be chosen to be other than for normal operation consistently with very important maintenance ,n

of radial conditions of the blade elementS., (In a turbine wheel receiving radially directed inliow the inlet angle consistent with radial blade conditions can only be 90.) In the case of thin blades in a diagonal flow wheel the inlet would theoretically be highly twisted in order to result in shockless entrance throughout the entire orthogonal extent of the entrance. The type of turbine here in disclosed avoids all these disturbances` Because of such a high tolerance to a substantial dii'erence between the blade angle -and the tlow angle, twist of the blading is not necessary' so that diagonal-flow bladngmay be used, Furthermore, the tolerance to the diierence of?r llow and blade angles broadens the high eiiciency region of the characteristic of eiiiciency plotted against rtl/cn ratios, as referred to above, to avoid the drops of ellis,

ciency at the high and low values of this ratio, with parf tisular avoidance of the braking action at high values of this ratio,

Rigidity and the withstanding ot centrifugal stresses has been mentioned, the bulbous structure at the inlet, resulting as a byproduct of the passage design, providing a very strong design in which the stresses are substantially solely in tension. Tied in with this is also blade rigidity resisting the building up of any vibration patterns. Vibration failure possible is made vanishingly small by the rigidity of the construction coupled with the prevention of flow separation which, if it occurred would set up transients serving as a source of vibration.

The blades are further resistant to damage by particles of solid nature carriedwith the stream. Thin edges of prior blades have been susceptible to such damage. With the present turbine construction particles directed at high velocity by the iow against the blades strike a smoothly rounded surface which is tightly stretched by centrifugal action and are readily carried away by the highly accelerated flow. Similar particles in prior constructions have involved repeated rebounding against the blade tips because llow velocities were insufficient to carry them through the ilow passages against centrigual forces.

FIGURES to 9 show a modification embodying the invention and involving inlet ow, without an axial component, an axial discharge flow. Provision is made in this construction for high effective cooling such as is required if the turbine is driven by high temperature cornbustion gases directly following combustion, or by higher temperature exhaust gases from a spark ignition engine, as contrasted with exhaust gases from a diesel engine, having relatively lower temperatures. The design illustrated in these figures is particularly applicable to a turbine having a rotor with an overall diameter of about two inches.

The figures considered jointly will indicate the threedimensional aspects of the rotor construction. In a turbine of this type introduction of gases to the rotor passages may be either as described above, involving a vortex ow inwardly, or nozzles may be used in conventional fashion to direct the gases. For exemplary purposes, one of the nozzles is indicated in FIGURE 5 at 140.

The hub of the turbine rotor is shown at 142 and is provided with the shaft extensions 144 and 146 which will be mounted in suitable bearings. The mountings and associated elements may be as already generally described with reference to FIGURE l.

The gas passages are indicated generally at 148, the entrance region being indicated at 156 and the discharge regions at 152. These passages are iirst designed in accordance with the principles already fully described, and the sections of these passages orthogonal to the llow vary along the flow path as set forth in detail heretofore, all of the discussed matters being here also involved and particularly those set forth in connection with the description relating to FIGURE 4 as to variation of the areas orthogonal to the flow with respect to mean line ilow distance.

When these principles of design are followed, it is again found that they must be bounded by surfaces giving rise to relatively thick blades. The shapes of the resulting blades will be apparent from the ligures, the blades involving leading surfaces indicated at 154 and trailing surfaces indicated at 156. rl`he inlet edges of these blades, which extend axially, are rounded and, in View of the hollow form of the inlet sections of the blades are provided by arches 157. The exit edges of the blades are shown at 158, and in FIGURE 6 the effective discharge angle is indicated at F, being illustrated as 37 at the diameter illustrated. The discharge in this modification has no radial component. The discharge may be followed by a diffuser as previously described.

vThe rotor illustrated has only six blades, and the scale atthe right of FIGURE 6 will serve to indicate in degrees the variations of the blade surfaces with respect to angles about the axis of rotation. The alignment of the section shown in FIGURES 6 and 7 may be noted from the line G in the latter which corresponds to 0 of the scale in FIGURE 6. The angular pitch is, of course, 60, and the angular scale on FIGURE 7 may be visualized by considering that the blade edges 158 are spaced by 60.

It may be noted that the trough of the gas passages is indicated at 16u, being in circumferential projection in FIGURE 5 and appearing in section in FIGURE 7. The trough in its entirety is in FIGURE 6, but appears only as the region between the blades.

For eective cooling advantage is taken of the fact that the blades are thick in the inlet region and-consequently may be hollow. The small diameter portion of the hub is provided with air inlet scoop passages at 162 which communicate with the radially extending passage portions 164 into the main hollow region 166 within each blade.

The regions 166 are bounded at their outermost radii by the arches 157, but are in communication with the surrounding housing axially. Discharge of air is effected through the nozzles provided at 168 delimited by the lips 170 provided in the radially outward regions of the leading surfaces of the blades. The discharged air merges with the driving gases in the form of sheets as will be obvious from consideration particularly of the upper portion of FIGURE 8.

No special remarks need be made concerning the operation of the last described modification since its characteristics and operations are identical with those previously described with respect to the lrst modication.

It will be evident that the turbine constructions which have been described are of general applicability. Various changes may be made in details without departing from the invention as defined in the following claims.

What is claimed is:

1. An elastic fluid turbine comprising a rotatable hub, blades extending substantially solely radially carried by the hub, a stationary housing bounding, with clearance, said blades, and means for directing elastic fluid with a substantial radially inward component of motion to said blades to drive Ythe same, said hub, blades and housing delining elastic fluid passages bounded, from their inlet ends to their outlet ends, radially inwardly by the hub, radially outwardly by the housing, and circumferentially by the blades, the mean radius of which passages is larger at their inlet ends than at their outlet ends, to constrain ow therein having, for at least the majority of streamlines thereof a radial component of dow, which component decreases from the inlet ends to the outlet ends of said passages, said blades having thick airfoil shapes providing for said lluid passages therebetween areas orthogonal to the :flow such that a plot of such orthogonal areas against a mean line distance of flow is concave upwardly, with continuous decrease of such orthogonal areas, thereby to provide a pressure gradient due to reaction which at each point of each passage exceeds the opposing pressure gradient due to centrifugal edects.

2.. A turbine according to claim 1 in which the blades are hollow and provide an entrance portion of arch-like form substantially free of bending stresses due to centrifugal forces.

5. A turbine according to claim 2 in which the hollow regions within the blades are open to the clearance between the blades and the housing.

4. A turbine according to claim 2 having passages for the ilow of cooling air into the hollow regions of the blades.

5. A turbine according to claim 1 in which said means for directing elastic iluid provides a whirl chamber of large radial extent in which inward vortex low is produced for approach to the blades.

6. A turbine according to claim 1 including a vaneless diffuser disposed to receive elastic liuid discharged from said passages, said diffuser extending substantially in a radial direction from its inlet portion to its outlet portion to convert kinetic energy of the gases discharged from said passages into a pressure rise.

7. A turbine according to claim 1 having passages for the ow of cooling air into the hollow regions of the blades, the last mentioned passages having impeller portions extending substantially radially outwardly to effect centrifugal action on the cooling air. 5

References Cited by the Examiner UNITED STATES PATENTS 1,777,098 9/ 1930 Lysholm 253-77 2,258,793 10/1941 New 253-39 10 2,407,531 9/ 1946 Bix-mann.

10 Birmann. Schneider. Weinberg 253-3915 X Birmann.

McDowall 253-39.15 Dundore et al. 60-54 Black et a1. 60-54 Brmann 230-114 X SAMUEL LEVINE, Primary Examiner. JULIUS E. WEST, Examiner. 

1. AN ELASTIC FLUID TURBINE COMPRISING A ROTATABLE HUB, BLADES EXTENDING SUBSTANTIALLY SOLELY RADIALLY CARRIED BY THE HUB, A STATIONARY HOUSING BOUNDING, WITH CLEARANCE, SAID BLADES, AND MEANS FOR DIRECTING ELASTIC FLUID WITH A SUBSTANTIAL RADIALLY INWARD COMPONENT OF MOTION TO SAID BLADES TO DRIVE THE SAME, SAID HUB, BLADES AND HOUSING DEFINING ELASTIC FLUID PASSAGES BOUNDED, FROM THEIR INLET ENDS TO THEIR OUTLET ENDS, RADIALLY INWARDLY BY THE HUB, RADIALLY OUTWARDLY BY THE HOUSING, AND CIRCUMFERENTIALLY BY THE BLADES, THE MEANS RADIUS OF WHICH PASSAGES IS LARGER AT THEIR INLET ENDS THAN AT THEIR OUTLET ENDS, TO CONSTRAIN FLOW THEREIN HAVING, FOR AT LEAST THE MAJORITY OF STREAMLINES THEREOF A RADIAL COMPONENT OF FLOW, WHICH COMPONENT DECREASES FROM THE INLET ENDS TO THE OUTLET ENDS OF SAID PASSAGES, SAID BLADES HAVING THICK AIRFOIL SHAPES PROVIDING FOR SAID FLUID PASSAGES THEREBETWEEN AREAS ORTHOGONAL TO THE FLOW SUCH THAT A PLOT OF SUCH ORTHOGONAL AREAS AGAINST A MEAN LINE DISTANCE OF FLOW IS CONCAVE UPWARDLY, WITH CONTINUOUS DECREASE OF SUCH ORTHOGONAL AREAS, WHEREBY TO PROVIDE A PRESSURE GRADIENT DUE TO REACTION WHICH AT EACH POINT OF EACH PASSAGE EXCEEDS THE OPPOSING PRESSURE GRADIENT DUE TO CENTRIFUGAL EFFECTS. 